Grinding machine



Aug. 26, 1941. c. HERFURTH ETAL 2,254,010

GRINDING MACHINE Filed April 25, 1938 9 Sheets-Sheet 1.

SEE/'03 7 ATTORNEY.

Aug. 2 194 c. HERFURTH Em 2, 4 0

GRINDING MACHINE Filed April 25, 1938 9 Sheets-Sheet 2 INVENTOR. 01 4/4245 dam/Pm fimvona. fl dam/raw: Wm

ATTORNEY Aug. 26, 1941.

C. HERFURTH ET AL GRINDING MACHINE Filed April 25, 1938 9 Sheets-Sheet 4 ATTORNEY.

Aug. 26, 1941.

C. HERFURTH ETAL GRINDING MACHINE I Filled April.25, 193a 9 Sheets-Sheet 5 1941- c. HERFURTH EI'AL 2,

GRINDING MACHINE Filed April 25, 1938 9 Sheets-Sheet 6 QHN ATTORNEY. V j

Aug. 26, 1941. c. HERFURTH ETAL 2,254,010 I GRINDING MACHINE 7 Filed April 25, 1938 v 9 Sheets-Sheet 7 I INVENT OR.

ATTORNEY.

26, 1941- c. HERFURTH r AL 2,254,010

' GRINDING MACHINE Filed Apfi l 25, 1938 9 Sheets-Sheet a M ATTORNEY.

Aug. 26, 1941. c. HERFURTH ETAL 2,254,010 7 GRINDING MACHINE Filed April 25, 1938 9 Sheets-Sheet 9 ATTORNEY.

Patented Aug. 26. 1941 GRINDING MACHINE Charles vHerfurth and Raymond D. Wortendyke,

Cincinnati, Ohio, assignors to Cincinnati Grinders Incorporated, Cincinnati, hio,-a corporation of. Ohio Application April 25, 1938, Serial No. 204,083

14 Claims.

machine which will execute an improved traverse grinding cycle wherebysudden changes in the rate of traversing movement will be eliminated so that a more uniform grinding operation may be obtained.

Another object of this invention is to provide a control mechanism for the traversing movement between a work piece anda grinding wheel, which has versatility as respects both adjustment and control, whereby variations in rate of any portion of the automatic cycle may be readily made, and selection between different types of cycles may be easily obtained.

A further object of this invention is to provide improved means for synchronizing the movement of a work support and a camber control mechanism.

An additional object of this invention is to provide in a hydraulic transmission, improved and in the specific structural details within the scope of the appended claims without departing from or exceeding the spirit of the invention.

Referring to the drawings in which like reference numerals indicate like or similar parts:

Figure 1 is a front elevation of a machine tool embodying the principles of this invention.

Figure 2 is a section on the line 2-2 of Figure 1 showing the gear train between the table motor and the table.

Figure 3 is a vertical section through the machine taken on the line 3-3 of Figure 1 showing part of the drive from the table to the cambering mechanism. I

Figure 4 is a horizontal section taken on the line 4-4 of Figure 1 showing the servo-motor control mechanism for the table motor.

Figure 5 is a section on the line 5-5 of Figure 4.

Figure 6 is a section on the line 6-6 of Figure 4;

Figure '7 is a section on the line l-l of Figure 4.

Figure 9 is a rear elevational view of the cambering attachment gear box as viewed on the line 9-9 of Figure 3.

Figure 10 is an expanded view of the gearing shown in Figure 9 and as viewed on the line Ill-III of that figure.

Figures 11, 11A, 12; 13 and 13A are views showing different settings of the camber control cam for obtaining difierent degrees of camber.

Figure 14 is a diagrammatic view of the hydraulic control for obtaining synchronism between the movement of the table and the movement of the camber control cam.

Figure 15 is a diagrammatic view of the hydraulic traverse transmission and control mechanism for the table.

Figure 16 is a detail section on the line Iii-l6 of Figure 4.

A machine embodying the principles of this invention is shown in Figure 1 of the drawings, and in general comprises a bed l0 upon the top of which are formed guideways H and I2 as more particularly shown in Figure 2. A work supporting table I3 is mounted on these guideways for traversing movement relative to a grinding wheel l4.

As shown in Figure 3, the grinding wheel I4v may be supported on a spindle l5 for rotation by 'a motor It. The grinding wheel and motor are mounted on a subslide H. which is movable relative to a support l8, which in turn is pivoted for movement about an axis l9 located under the grinding wheel. I

Suitable mechanism may be employed for effecting movement of the slide I1, but since this mechanism forms no part of the present invention, it has not been illustrated. Referring to Figures 2 and 4, the mechanism for traversing the table comprises a rotary hydraulic motor 20 which is operatively connected to a drive shaft 2| for rotating a worm 22 antifrictionally supported in the bed. This worm meshes with a worm .wheel 23 secured to the lower end of shaft 24 which is also antifrictionally supported in the bed. The upper end of this shaft has an integral pinion 25 meshing with a rack bar 26 secured to the under side of the table between the guideways II and i2. It will now be apparent that upon rotation of the hydraulic motor in one direction orthe other that the table may be traversed relative to the. grinding wheel. g I

The table l3 has a headstock 21, and a tailstock 28, mounted thereon for adjustment to ac- Figure 8 is a section on the line 8-8 of Figure 4. commodate difierent lengths of Work pieces. The

axis 29 which connects the points of the work supporting centers 30 and 3| may be defined as the rotational axis of the work. This machine is especially suitable for grinding rolls and is provided with an adjustable camber control mechanism which operates to oscillate the support I3 about the axis I9. Since this axis is under the grinding wheel, the lever arm which connects this axis to the center 32 of the grinding wheel moves practically at right angles to a line 33 which connects the axis of rotation29 to the center of the grinding wheel so that for small movements of the grinding wheel sufficient to produce the necessary camber on the work, it may be said that the grinding wheel moves toward and from the work along substantially a radial line thereof..

The mechanism for oscillating the support l6 to produce thecambering action consists of a camber control mechanism which is driven from the table rack 26 through a synchronized transmission which maintains the camber control mechanism in exact step with the movement of the table. The adjustable camber control mechanism to be described does not constitute part of the present invention, but is described and claimed in a separate copending application, Serial Number 204,082, filed April25', 1938. The rack 26 engages a pinion 34 which is keyed to the upper end of the shaft 36. This shaft is connected through a bevel pinion 36 and a bevel gear 31 to a longitudinal shaft". This shaft has a It will be noted that the pinion 62 has a reduced portion 65 which passes through a hole 66 formed in the clamping ring 61, and since the pinion 62 is fixed with the gear 49, the clamping ring 51 will also be held with the gear, whereby the cam 55'may be rotated relative to both of these parts.

As shown in Figures 3 and 9, a cam follower 61 is supported for vertical reciprocation in the bed It], and one end of this follower is provided with a roller 63 which engages the periphery of the cam 66, and the other end engages a wear plate 69 mounted on the under side of the pivoted support I8. It will be noted from Figure 9 that the vertical axis I6 defines a vertical plane in which the follower reciprocates, and that the axis passes vtive positions of the eccentrics for set-up purposes, a plate 12 is attached to the end of the splined end 39 fitting in a splined bore 46 integral with the bevelgear 31 so as to facilitate assembly of the parts. Referring to Figure 10, the shaft 36 is antifrictionally supported by bearings 4| in the wall 42 of the cambering mechanism gear box indicated generally by the reference numeral 43. The shaft 38 has a spur gear 44 keyed to the end thereof and intermeshing with a gear 46 keyed to a stub shaft 46. The gear 46 is held in position on the end of shaft 46 by a nut 41 threaded on the end of shaft. The gears 44 and 46 are removable and gears of different sizes may be substituted therefor whereby they constitute change gears for varying the rate of actuation of the remainder of the gear train.

The shaft 46 also has a gear 46 which intermeshes with a large gear 49. The gear 49 is supported by antifriction bearings 66 on a fixed shaft 5|. The gear 49 has an elongated eccentric hub 62 having a tapered portion 63 and a straight portion 54. A circular cam member 66, having an eccentric bore 66, is mounted on the hub of the gear 49. This eccentric bore has a straight portion and a tapered portion whereby the cam may be forced on the tapered portionand fricthreaded in the cam and have reduced ends 66 which engage the face 6| of the gear 49 whereby the reaction on the screw threads will, force the cam ofi of the taper 63. When the cam has been loosened in this manner, a small pinion 62 which is journaled in the gear member 49 by means of an integral stud 63, is rotated, and th'rou'ghinter- I engagement with internal gear teeth 64 carried by the cam member, the same may be rotated relative to the gear 49.

fixed shaft 5| by means of screws I3, and this plate is provided with two diametrically opposite index marks I4 and I5 that lie in a plane at right angles to the vertical axis I6. Referring to Figure 11 in which the eccentricities are exaggerated for explanatory purposes, the point 16 indicates the true center of the hub 52, and therefore the distance between the axis of rotation II and the point I6 represents the effective radius or eccentricity. When the cam 66 is rotated relative to the hub 62, the point I6 becomes the axis of rotation about which the cam 55 moves. When the cam is in the position shown in Figure 11, the true'center of the cam lies on the axis of rotation 'II because its eccentricity, which may be measured to the right of the point I6, is equal to and The result is that the true center of the cam lies on the axis of rotation, and the resultant eccentricity is equal to zero. Therefore with the parts in a zero-eccentric position, the end face of the hub 62 is graduated, beginning with a 0 mark opposite the arrow I4, and other marks extending throughout the arc clockwise to the axis I0. 6

Since the cam 55 moves about the center I6 when it is adjusted relative to the hub 52, it is provided with an arrow II which is located in the plane of the radius 13 passing through the center I6 and parallel to the vertical axis III. The end face of the hub 62 is graduated from this point through an arc of in a clockwise direction, the graduations being equal in number to the first named graduations but having twice the angular extent.

The graduations shown indicate percentages of the maximum camber which can be obtained.

. This maximum camber will be equal to twice the longitudinal center of the work as represented by the axis I9, Figure 14, is opposite the center of the grinding wheel I4, and then the gear 46 is withdrawn from mesh with the gear 44, so

that the gear 49 may be rotated in a counterclockwise direction, as viewed in Figure 11, until the desired camber percentage is opposite the arrow I4.

The gear 45 is then replaced in mesh with the gear 44, and the cam 55 loosened and rotated relative to the gear 49 until the arrow 11 is op- Dosite a graduation mark having a value equal to the value of the graduation mark opposite the arrow I4. The effect of this is to shift the true center of the cam circle 55 along the axis I0, and, in this case, downward or below the center 1| which will thereby cause the grinding wheel support I8 tomove clockwise as viewed in Figure 3, thus determining the maximum diameter or crown of the Work at its longitudinal center. It

will now be evident that after the cam 55 is reimum camber obtainable. To obtain this set-u the gear and cam are rotated about the axis II in a counterclockwise direction through an an le indicated by the reference numeral 80 until the graduation mark is opposite the arrow I4. The parts will thenbe in the position indicated in Figure 11A. It will be noted that the true center of the cam still remains in juxtaposition to the center II. The center I6 of the eccentric hub 52 has moved and lies on the radial line 8|. The cam 55 is now rotated about the center '76 and through an angle equal to twice the angle 80 so that the true center of cam 55 will lie on the vertical line I'll.

It is desirable that the maximum offset or eccentricity be measurable along the line I0, because this position of the cam corresponds to the central position of the work with respect to the grinding wheel and whatever maximum displacement the cam can effect should be effected while the work is central of the grinding wheel. If a radial line 82 is drawn from the cent r I6 parallel to the diameter through the arrows 74-75, the 'angle 83 between this line and the prolongation of radial line 8|, will be equal to the angle 80. If the angle 84 is made equal to the angle 83, the true center 85 of the circle 55 will lie on the axis 10 and the eccentricity is then the distance between the line 80, which was tangent to the circle 55 in Figure 11 and the new tangent 81 shown in Figure l2. this distance being measured along the axis I0.

The above set-up was for obtaining convex cambers on the work. To set up for concave members, the parts shown in Figure 1 are r tated through 180 so that the inner 0 mark on the hub 52 is opposite the arrow I5. as shown in Figure 13. The gear and cam will then rota together in a counterclockwise. direction unt l the proper graduation mark is opposite the arrow 15, and then the cam is rotated relative to the gear until the arrow 17 is opposite the same graduation mark on the outer periphery of t e hub 52. An example of such a set-up is shown in Figure 13A.

Since the weight of the grinding wheel sup port l8 rests on the follower pin .61, it is of course desirable that this weight be removed in order to facilitate adjustment of the cam during set-up. Power operated means have been provided for lifting the support I8 a suflicient distance that no weight will be on the pin 61 dur-v ing adjustment. This mechanism comprises a p ston 88, Figures 3 and 14, which is reciprocably mounted in a cylinder 89 formed in the bed I0. The piston has a reduced end 90 which passes through the cylinder head 9| for engagement with the plate 69. Hydraulic pressure is admite ted to port 92 which is in constant communication with an annular groove 93 formed on the periphery of the piston. An interdrilled passage 94 connects this groove to the bottom of the cylinder whereby the fluid pressure will act on the under side of the piston to raise the same. The shoulder 95 on the upper end of the piston limits its outward movement and determines the amount of lift that will be imparted to the support l8.

The admission of fluid pressure to the cylinder 89 is controlled automatically by a valve plunger 96 which is reciprocably mounted in a valve housing 91. A spring 98 is interposed in the end of the valve housing for continuously urging the valve plunger in one direction. The valve housing is contained within the gear box 43 and this gear box is provided with a removable cover. 99. When the cover is closed, it engages the end of the valve plunger 96 and moves the same inwardly into the position shown in Figure 14. The valve housing has a port I00 which is connected by a channel I 0| to the port 92 of the cylinder. It is also provided with a port I02 to which fluid is supplied through a channel I03 from a suitable source of pressure, such as a supply pump I'04. When the valve plunger is in the position shown, a spool I05 closes the port I02 thereby disconnecting it from the port I00. When the door 99 of the gear box is opened, the spring 98 moves the plunger outward, and an annular groove I00 formed in the plunger interconnects the port I02 with port I00 whereby the fiuid pressure flows to the cylinder 89 and causes outward movement of the piston 88. When the door is closed. the annular groove I06 interconnects port I00 with an exhaust port I01 whereby the fluid in the cylinder is exhausted to the reservoir I08.

As previously mentioned, it is desirable that the camber control cam rotate in exact synchronism with the table movement and this means that there should be no lost motion in the gearing between the table and the cam becafuse otherwise the maximum point of camber would shift for each direction of movement of the table. Means have therefore been provided for taking the backlash orlost motion out of the drive connection between the table and the cam, and this means or mechanism acts in such a manner that the backlash is always taken out in one direction only and regardless of thedirection of movement of the table. This is necessary in order to prevent shifting of the maximum point of camber relative to the work.

The mechanism for accomplishing this comprises an hydraulic motor I09, which is mounted in the gear box, as shown in Figure 10. The drive shaft H0 of this motor is provided with a spur gear III, which is connected by the double gear II2 to the gear 49. It will now be apparent that when the shaft 38 is rotated to cause actuation of the gear 49, that the gear 49 in turn will drive or tend to drive the gear I I I on the end of the motor shaft H0. In order that the backlash may always be taken up in one direction regardless of the direction of movement of the table, or, in

other words, the direction of rotation of theduring rotation of the shaft '38 in one direction,

the motor will act as a retarding or braking means, and during rotation of the shaft 38 in the other direction, the pressure in the motor will be increased so that the motor will serve as a driving means. Although the motor I09 would otherwise be able to do the driving and feed the table, it must be remembered that the feed motor 20 is connected to the table through a worm and worm gear drive and therefore the actual rate of movement of the table will still be determined by the fed motor so that the motor I09 although actually moving the table will do so at a rate determined by the rate of rotation of the feed motor 20.

This automatic change in the unidirectional pressure acting on the motor I09 is automatically eifected when the work supporting table is reversed and is effected by the automatic reversingv mechanism.

,The two diflerent pressures needed ion the operation of motor I09 are provided by the pump I04, and a selector valve H3 is provided for alternately connecting these pressures to the motor. The higher pressure is provided by connecting the output of pump I04 directly to port 4 of the selector ,valve. The lower pressure is obtained by connecting a branch line H5 in series with a pair of relief valves H6 and H1.

' The valve H6 is setzto yield the higher pressure when the valve plungers 96 and I24 are in the position shown in Figure I4, the selection of pressure delivered to the motor I 09 is determined by the selector valve I I3 and to this end the valve has a plunger I35 in which are formed two annular grooves I36 and I31. When the plunger I35 is shifted to the left, the annular groove I31 interconnects the lower pressure port I I9 to port I20, and when the plunger is shifted to the right.

The selector, valve II 3 has a port I20 which is connected by a channel I 2I to port I22 of a cut out valve I23 for the motor I 09. This valve has a plunger I24 in which is formed an annular groove I25 which, in the position shown, serves to' connect the port I22 to port I26. The port I26 is permanently connected to port I 21 of valve 91. The valve plunger 96 has an annular groove I28, which is in such a position, when the door of the gear box is closed, to interconnect port I21 to port I29 and thereby to channel I30 which leads to the hydraulic motor I09. The other port of the motor is permanently connected through channel I3I to the reservoir I08. When- I ever the valve plunger I24 is pulledout, a spool I32 closes the pressure port I22 and'the annular groove I25 interconnects port I26: to an exhaust port I33. Thus when the gear box is closed, and the valve plunger I24 is pulled out, both ports of the motor I09 are connected to reservoir thereby offering no resistance or braking action on the table drive mechanism. When the valve plunger I24 is in the position shown in Figure 14, and the gear box cover is opened, the shifting of valve plunger 96 will disconnect port I26 from port I29, and the annular groove I28 will interconnect the port I 29 to the exhaust channel I34 whereby the pressure will be oif of the motor I09, so that the gear 49 may be rotated readily for set-up purposes.

the annular groove I36 interconnects the higher pressure port II4 to port I20.

The plunger I is shifted automatically in accordance with the direction of movement of the table, and the ends of the valve housing H3. are provided with ports I38 and I39 which are connected respectively through channels I40 and I to a pilot valve I42, shown in Figure15. The pilot valve I42 has pressure ports I43 which are connected by a channel I44 to a suitable source of operating pressure; and ports I45 and I46 on opposite sides thereof to which are connected the channels I and I40 respectively. Thevalve plunger I41 is self actuating in that it has a central spool I48 which is larger in diameter than the two end spools I40 and I50.

When the plunger is to the left, the annular groove I5I interconnects pressure port I43 to port I46 to cause shifting of the plunger I35 to the right. A trip operable mechanism is provided for shifting the plunger I41, but it will be noted that after the plunger passes the pressure port I43 in either direction, the pressure differential created by the difference in area between the spool I48 and the end spools will complete the shifting movement thereby simulating the action of a conventional detent mechanism.

The trip mechanism comprises a trip plunger I53 from which projects a pair of wings I54 and I55, the wings lying In diflferent planes, as shown in Figure 1. These wings are alternately engaged by dogs I56 and I51 carried by the table I3 for rotating the plunger. The plunger has a ball ended lever I58 which engages a groove -I59 formed in the valve plunger I41 for shifting the same. It will thus be evident that the direction of movement of the table determines the position of the selector valve plunger I35 and thereby the value of the unidirectional pressure to be applied to the motor I09.

The trip plunger not only determines the pressure to be applied to the backlash motor, but

also controls the reversing of the feed motor 20, and in such manner as to cause deceleration of the motor at a uniform rate prior to the shifting of the reverse control valve. The entire cycle of reversing, however, includes not only the deceleration of the'table, but also means for effecting a predetermined dwell of the table, and a controlled uniform acceleration in the opposite direction. This manner of reversing the table eliminates any sudden changes in rate which might cause'undesirable vibrations and their resultant eifect upon the finish of the work. These various results are obtained by a reversing control valve I00, which has an acceleration and deceleration control valve plunger I6I, and areversing control sleeve I62. The plunger I'6 I is connected by a ball ended crank I63 to the trip plunger I53 for actuation thereby.

The reversing valve I has a pair of motorports I64 and I65, which are connected by channing position shown, interconnects ports I69 and I68 to ports I12 and I13 and thereby through channels I14 and I15 to the motor 20.

The valve I60 also has a pair of pressure ports I16 and "1 which are supplied through a channel I18 from port I19 of a rate control valve indicated generally by the reference numeral I80. The valve also has a pair of exhaust ports I8I and I82 which are connected to the return channel I83. The reversing valve. sleeve I62 has a series of annular grooves I84, I85, I86, I81, I88, and I89 which are in continuous communication with ports I8I, I64, I16, I11, I65, and I'82 respectively. Each of these annular grooves has a series of radial holes drilled in the bottom thereof and communicating with the interior of the sleeve, thus forming ports for control by the inner plunger.

In the-position of the parts shown in Figure 15,

the table I3 has been moved toward the left and the beveled face I90 on the dog I51 has engaged the curved face I9I on the wing I55 and rotated the plunger I53 in a counterclockwise direction, and this rotation has taken place through a predetermined length of movement of the table.

- 20 and thereby of the table I'3.

The final shifting of the pilot valve caused admission of pressure to port I46 and thereby to branch channel I92 leading to port I93 located in the left hand end of valve housing I60. This resulted in pressure being admitted to cause shifting of the reversing sleeve I 62. The space I94 in the other end of the valve housing is in communication through an interdrilled passage I95 in the sleeve to port I96 and this port is connected by channel I91 to ports I98 and I99 of a combination tarry and acceleration control valve 200. The fluid entering port I99 of this valve passes through a spring closed check valve 20I to an interdrilled passage 202 which terminates in an annular groove 203 formed in the tarry control valve plunger 204. A spool 205 on one side channel 2 to port 2I2 of a combination tarry and acceleration control valve indicated generally by the reference numeral 2 I3. This port communicates with a T-shaped interdrilled passage 2| 4 which terminates in a spring pressed ball check valve M5. The fluid flows past this check valve into a spiral groove 2I6 which has sulficiently small pitch to offer a definite resistance to the flow of fluid through it.

This spiral groove terminates in a port 2I I which is connected by channel 2I8 to channel 208 leading to port I45 of the pilot valve I42. The port I45 is connected at this time to an exhaust port 2I9 through which the fluid flows to the return channel I83. of the reversing valve sleeve I62 will depend, upon first the resistance of the tarry control grooves 206, which may be different from the resistance of the acceleration control spiral groove 2I6. Therefore a first means has been provided for determining the rate of the first half of the stroke of the reversing valve sleeve to determine the amount of tarry of the table, and a second means for determining'the rate of the last part of the stroke of the reversing valve sleeve to determine the rate of acceleration of the table in the new direction.

After the pressure port I11 has become fully connected to port I65, the rate of fluid supply to the motor 20 is determined by the rate control valve I80. The supply of operating fluid for the motor 20 comes from a separate pump 220 through channel 22I to port 222 of a combining The rate valve has an adjustable plunger 228.

which is urged in one direction by a spring 229 to close port 221, and in the other direction by a separate plunger 230 which is threaded in the bed of the machine. The plunger has a series of longitudinally extending V-shaped grooves 23I which decrease in depth at one end so that as the plunger 228 is moved against the resistance of spring 229, the resistance of port 221 may be gradually reduced. In this manner the flow of of this groove has a series of v shaped grooves 206 formed in its periphery which control the rate of flow to a port 201. The port 201 is connected by channel 208 to port I45 of the pilot valve I42, and at this time this port is connected to the return channel I83. Thus the rate of movement of the sleeve, I62 toward the right depends upon the setting of the tarry control valve plunger 204. This will determine the length of tarry of the table, which will continue until the sleeve I62 is shifted toward the right, a sufllcient distance to interconnect the pressure port I11 with the motorport I65. I

, At this point in the cycle of movement of the sleeve toward the right, the port I96 will close and a second port 209 will just start to open. In

other words, the end of the sleeve is provided with a second interdrilled passage 2I0 which will now be placed in communication with the port 209. This means that the fluid that is being exhausted from the space I94 will stop flowingv through port,

I96 and will start to flow through port 209 and fluid to the hydraulic motor 20 is throttled to yield various feed rates.

When therate valve is adjusted to a position to yield the higher feed rates, the pump 220 will not deliver sufficient volume to produce these rates, and therefore an auxiliary pump has been provided together with suitable control means for automatically combining or adding the delivery of the auxiliary pump to that of the main pump. In order, however,-to prevent the auxiliary pump from laboring when not being utilized, a by-pass connection' has been provided and this by-pass control is so arranged that'the discharge'from the auxiliary pump is gradually throttled asthe' setting of the rate valve goes up, so that desirable.

I division of flow of theoutput of the auxiliary pump is made in accordance with the demands for extra fluid and in accordance with the degree of increase in the rate.

As shown in Figure 15, the auxiliary pump 232 is connected'by a channel 233 to port 234 of'the "combiningvalve 223. The port'234 is normally maintained closed by a'spring pressed plunger 235 and the I spring is .ass isted by. the pressure existing in channel 226. there being a branch connection 236 from thischannel to theend of the bore 231 which contlains the plunger The Thus the rate of movement periphery of two discs 269 and 264. As shown in pressure that will exist in the channel 239 tending to shift the valve plunger 285, will depend upon the setting of an auxiliary valve plunger 238 which is mounted in the same block with the rate valve and adjusted by the same control. This valve has a port 239 to which a branch 248 of channel 233 is connected, and when the rate valve is set for low feed rates, the port 239 is connected to a reservoir port 2 by means of am annular groove 242 formed in the plunger 238.

When the plunger 230 of the rate valve isadjusted in a direction to increase the feed rate, a laterally extending arm 243 which is connected to the plunger 238 moves this plunger simultaneously with plunger 228 and in a direction to gradually close the port 242 and thereby increase the pressure in channel 240 to such a point that when the pressure in the channel 224 drops below a predetermined minimum, the valve 285 will open and thereby combine the delivery of the auxiliary pump with that of the main pump. Automatic reciprocation of the table is stopped by shifting the plunger l1I of the start and stop control valve downward, as viewed in Figure 15,

by rotating the control handle 244 to a central position. This will close ports I69 and I68 thereby stopping the fiow to and from the motor. The table movement control mechanism just described is specifically claimed in our copending application, Serial No. 372,240.-

When it is desired to control the movement of the table manually, the lever 244 is rotated to the leftof its central position thereby depressing the valve plunger l1 l a suflicient amount to interconnect the motor port I 12 with port 245, and the motor port I13 with port 246. This serves to interconnect the hydraulic motor 28 with a servo control valve 241 through channels 248 and 249 which terminate in ports and 25I of the servo-valve. The servo-valve also has a pressure port 252 and an exhaust port 259. The pressure port 252 is connected bychannel 254 to port 255 of an auxiliary control valve 256 which has a plunger 251 which is simultaneously reciprocated with the plunger I1l under control of lever 244, but in an opposite direction. Therefore, the plunger 251 moves up when the plunger I1l moves down, and this results in port 255 being interconnected by' the annular groove 258 to port 259. The port 259 is connected by a branch channel to the main pump supply channel I44. Therefore,

Figure 4, the disc 263 is fixed on the drive shaft 2I for rotation by the hydraulic motor 29. The disc'264 is supported on the end of a shaft 266 which is antifrictionally supported in the bed for rotation and is operatively connected to the rotary plunger 266 of the servo-valve.

There is a gear 261 which is sup for rotation on a reduced diameter 268 of the servo-valve housing and this gear is connected by a series of pins 269to an equalizer plunger housing 219 supported on the end of the rotary plunger 266 for movement relative thereto but connected as by pins 2" to the rotary valve sleeve 212.

"In other words, when the sleeve 212 is rotated by the gear 261', the motor 20 is caused to rotate and through the teed-back connection comprising the disc 263, the roller 262 and the disc 264, the

. plunger 266 is caused to rotate and thereby follow up the movement of the sleeve in a direction to stop operation of the motor 20.

The gear 261, as shown in Figure 1, is con-' nected by a gear train indicated generally by the reference numeral 213 to gear 214 secured to shaft 215 which carries a hand wheel 216. Thus, by rotation of the hand wheel 216, the servo-valve sleeve may be rotated to cause operation of the hydraulic motor 28 in opposite directions.

The manner in which the hydraulic connections to the motor are eflected by the servo-valve will now be explained. The servo-valve plunger 266 has a series of longitudinally cut away portions whereby a cross section through the plunger resembles a quadrate cross. This forms tour quadrants 211,218, 219, and 288. As shown in Figure 5, the pressure groove 252 has a pair of radial ports 28l and 282 which are normally closed when the valve is in a stop position. From this it will be seen that when the sleeve 212 is rotated in a counterclockwise direction relative to the plunger 266, that the quadrants 211 and 219 will interconnect the radial pressure ports 28I and 282 with radial ports 289 and 284 in the groove 250 which leads to one port of the motor. At the same time the radial ports 285 and 286 in groove 25l, as shown in-Figure 7, will be con.- nected to the radial ports 281 and 288 in exhaust groove 253, as shown in Figure 8. Thus the other central position, the pressure port 259 will be closed, but upon further downward movement into the position shown in Figure 15, the pressure port 259 will be connected to a channel 289 leading to the other end of cylinder 268 whereby the plunger 26l will be moved upward and thereby break the feed-back connection from the motor 20 to the servo-valve. In order to prevent the plunger 266 andthe sleeve 212 from getting out of phase during automatic operation of the machine, the equalizer plunger-housing 219 is pro- "vided with twodiametrically opposed equalizing plungers 299 and 29l, Figure 16, which have V- shaped ends fitting inJV-notches 292 and 299 formed in opposite sides of a collar 294 fixed with the rotary plunger 269.

\ When the table I9 moves toward the right. as

viewed in Figure 15, the inclined surface 296 on the dog I66 engages the wing I64 of trip plunger I69 rotating the same in a clockwise direction, thereby shitting the, pilot valve control plunger to the right and. aimultaneouslymovingthe decelerati on control valve plunger I6I in the same direction. The result is the gradual closing of p I e port I11, whereby the rate of movement of the table will decrease.

The fluid pressure that is admitted to port I46 by the shifting of the pilot valve will flow through channel 269 to port 296 of control valve 2I9.

This valve has a throttle valve plunger 291 similar to valve plunger 205 which throttles the flow from port 296 into the annular groove 298. The fluid continues through an interdrilled passage 299 and past spring closed check valve 300 to channel 2 through which it continues to the right hand end of valve I60. This will cause movement of the sleeve I62 toward the left until the port 209 closes.

Since the fluid can no longer flow through valve 2I3 from channel 208 on account of the closing of port 209, it will be forced through port 30I and spiral groove 302 of valve 200, to the int'erdrilled passage 303, which terminates in a spring closed check valve 304. This valve will open, and the fluid will continue through port I98 and channel I91 to port I96 of valve I60, thereby continuing the movement of the sleeve I62 to the left and causing the motor to accelerate in the new direction. It is thus possible to cause the table to decelerate to a stop, tarry, and accelerate in a new direction at each end of the table stroke.

There has thus been provided an improved transmission and control mechanism for a grinding machine which is capable of yielding an improved traverse cycle; which is adjustable to vary the feed rate and the length of tarry; which is provided with a camber control mechanism that is accurately synchronized with the table movement and adjustable to yield variable cambers; and which is provided with a manual control mechanism having intermediate servo-motor actuating mechanism, whereby manual operation and/or automatic cyclic operation, with or without a cambering mechanism, may be obtained.

What is claimed is:

1. In a machine tool having a work support and a tool support, the combination of transmission means for effecting relative movement between the supports, including a first power driven train for moving one of said supports, a second train in driving engagement with the movable support and including an actuator for the second support, and means to eliminate backlash from said second train to maintain synchronism between said actuator and the movable support, including a hydraulic motor connected to the second train, and means for maintaining unidirectional pressure on said motor whereby the to said work support, a reversible hydraulic motor i connected by a motion transmitting train to said rack to control the reciprocation of the work support, a control cam for shifting the grinding wheel toward and from the work to camber the same, a motion transmitting train coupled with said rack driving connections between said cam and train for rotating said cam, said last named train terminating in a hydraulic motor, and means to apply a variable unidirectional hydraulic pressure on said motor whereby during one direction of movementof the work support the motor will act as a retarding means, and during the other direction of movement the motor will serve as a driving means. 3. In a grinding machine having a work support and a grinding wheel support, the combina-' tion of means for reciprocating one of said supports and oscillating the other support in synchronism therewith, including a pair of hydraulic motors, a motion transmitting train extending from one of said motors to the other, serially arranged means in said train for actuating each of said supports, means for applying a uniform hydraulic pressure to one of said motors, and a relatively lower pressure to the other of said motors to effect actuation of the train in one direction, and means to reverse said one motor and increase the pressure on the other motor to effect actuation of the train in an opposite direction. i

4. In a grinding machine having a work support and a grinding wheel support, the combination of means for reciprocating one of said supports and oscillating the other support in synchrom'sm therewith, including a pair of hydraulic motors, a motion transmitting train extending from one of said motors to the other, serially arranged means in said train for actuating each of said supports, means for applying a uniform hydraulic pressure to one of said motors, and a relatively lower pressure to the other of said motors to effect actuation of the train in one direction, and means to reverse said one motor and increase the pressure on the other motor to effect actuation of the train in an opposite direction.

5. In a grinding machine having a reciprocating work support and a grinding wheel movable toward and from said support the combination of a reversible hydraulic motor, a motion transmitting train connecting said motor to the work support, a unidirectionally driven motor, a train extending from the work support to the last-named motor and having means therein for oscillating the grinding wheel, a valve for reversing the fluid pressure on the first named motor, a pressure selector valve for the second named motor, and a common pilot control valve for shifting said reverser and selector valves.

6. In a machine tool having a relatively mov-' able tool support and work support, the combination with a reversible power operable motor for determining the rate of said relative movement in opposite direction, of fluid operable means connected to the moving support for opposing one direction of movement and assisting the other direction of movement thereof, a pressure determining valve havir'ig two positions for connecting fluid supplies of different pressure to said means, valve means for controlling said motor, including a valve member having movement through a neutral zone for creating a tarry in said motor, and a pilot valve having fluid connections for simultaneously shifting said pressure determining'valve and said valve member.

7. In a machine tool having a relatively movable tool and work support, the combination of a hydraulic control circuit for effecting said movement in opposite directions, including a pair of fluid operable driving motors, a two-pressure system forselective actuation of one of said motors, including a shiftable pressure selector valve, a decelerator valve member and a tarry valve member for controlling said driving motors, a pilot valve, a trip plunger having operative 'connection for simultaneously shifting said pilot valve and decelerator valve, fluid operable means for completing the movement of said trip plunger and thereby of the connected members, and fluid "pressure means connectible by said pilot valve for shifting of said pressure selector valve and of said tarry control valve member.

8. In a machine tool having a work support and a tool support mounted for relative movement in two angularly related directions, means for controlling the relative movements of said supports including a first motor driven train connected to one of said supports, an actuating motor therefore, means for reversing the directional actuation of the motor, a second train coupled to the motor for transmitting motion to the second support in synchronized relation to the movement of the first support, and hydraulic means for eliminating backlash in the driving trains for the supportsincluding a unidirectionally actuated hydraulic device reacting on said trains in opposition to the motor.

9. In a machine of the character described including a shiftable part, a pair of simultaneously power actuated hydraulic motors coupleable with the part in opposed driving relation for eiiectinga shifting movement thereof, means for determining-the rate of actuation of one of said motors, and additional means for determining the relative power reaction of the motors whereby they selectively act as drivers during joint power actuation for movement of the part at the determined rate.

10. In a machine of the character described including a shiftable part, a pair of simultaneously power actuated hydraulic motors coupleable with the part in opposed driving relation for effecting a shifting movement thereof, means for determining the rate of actuation of one of said motors, and additional means for determining the relative power reaction of the motors whereby they selectively act as drivers during joint power actuation for movement of the part at the determined rate, said means including an element activated by movement of the part for effecting I reversal of said movement.

11. In a machine of the character described including a shiftable part, a power source a pair of hydraulic motors simultaneously coupleable 12. In a machine of the character described including a shiftable part, a pair of hydraulic motors simultaneously coupleable with the part for effecting a shifting movement thereof, means for determining the rate of actuation of one of said motors, additional means for determining the relative driving reaction of the motors for movement of the part at the determined rate, said means including an element activated by movement of the part for effecting reversal of said movement, a direction determinator for one of said motors, and an activating pressure deter.-

minator for the other of said motors actuable by said element.

13. In a machine of the character described including a shiftable part, a pair of hydraulic motors simultaneously coupleable with the part for effecting a shifting movement thereof, means .for determining the rate of actuation of one of said motors, additional means for determining the relative driving reaction of the motors for movement of the part at the determined rate, said means including an element activated by movement of the part for effecting reversal of said movement, a direction determinator for one of said motors, and an activating pressure determinator for the other of said motors actuable by said element, one of said determinators including hydraulically operable means for completing the actuation of the element and associate parts as initiated by movement of the shiftable part.

14. In a machine of the character described including a shifta-ble part, a pair of hydraulic motors simultaneously coupleable with the part for'eflecting a shifting movement thereof, means for determining the rate of actuation of one of said motors, additional means for determining the relative driving reaction of the motors for movement of the part at the determined rate, said means including an element activated by movement of the part for efiecting reversal of said movement, a direction determinator for one of said motors and an activating pressure determinator for the other of said motors actuable by said element, both of said determinators including hydraulically operable means for completing the actuation of the element and associate parts as initiated by movement of the shiftable part. I

CHARLES I-IERFUR'I'H. RAYMOND D. WORTENDYKE. 

